Actuator for transfer case

ABSTRACT

A transfer case for a vehicle having two output shafts, a gear reduction assembly, a coupling mechanism and an overrunning roller clutch for selectively producing driving of one shaft only or both shafts concurrently. The coupling mechanism selectively couples one output shaft to either (1) an input shaft, (2) the gear reduction assembly, or (3) a neutral position. The overrunning clutch has an inner race, an outer race, and rollers located between the races. Drag shoes are positioned to frictionally slide on a drag surface of a selectively grounded member to retard the rollers. A resilient band urges the drag shoes against the drag surface. When the drag shoes rotate at a sufficient speed they disengage from the drag surface so as to provide no force to retard the rollers. When the ground member is grounded it provides the drag surface for the drag shoes. When the ground member is ungrounded it is free to rotate and the drag shoes do not provide a drag force to bias the rollers. A latch may be coupled to the inner race to engage a roller cage to prevent high speed lock-up. A drag ring is located inside the outer race and provides a drag force on the rollers to advance the rollers when it is desired, e.g., when front wheel compression braking is advantageous. An actuator assembly is provided with a variable speed drive for shifting the transfer case between modes of operation.

BACKGROUND OF THE INVENTION

1. Field of the Invention

This invention pertains to transfer cases in vehicles having multipledrive modes, and more particularly pertains to transfer cases having anactuator for shifting between drive modes.

2. Description of the Related Art

Four-wheel drive vehicles generally incorporate some manner of transfercase by which torque from a single output shaft from a power source istransferred to two output shafts for driving separate axles of avehicle. In a standard configuration of a four-wheel drive vehicle therear wheels constantly receive torque from the power source and thefront wheels receive torque selectively, for example, "on-demand," whenthe rear wheels slip, or "part-time," when an operator shifts thetransfer case to four-wheel drive mode.

Various transfer case designs address different vehicle operatingconditions and requirements. One style of transfer case incorporates aninput shaft having a splined end, and a fixedly attached sun gear thatcooperates with a planetary gear assembly. A shifting mechanism couplesone output shaft to the planetary gear assembly or directly to the inputshaft to provide different ranges of operation (e.g., low range, highrange). That one output shaft may then be selectively coupled to asecond output shaft to transfer torque thereto.

One method of transferring torque between output shafts uses anoverrunning roller clutch. Such a transfer case is shown in FogelbergU.S. Pat. No. 4,124,085. Basic roller clutch design is known in the art.

The basic design of a roller clutch, without the overrunning feature,has concentric races with rollers (preferably needle rollers) locatedbetween the races. One of the races (typically the inner race) has aplurality of cam surfaces for engaging the rollers and thus isdesignated the driving race or driving member. When either race rotatesfaster than the other, the rollers jam (i.e., lock) the cams and thedriven race, thus engaging the clutch to transfer torque. Each rollercan lock in one of two positions: a "retarded" position, that is at atrailing edge of a cam surface or an "advanced" position located at aleading edge of each cam surface.

In an overrunning clutch, the rollers are biased into a retardedposition, that is biased opposite the direction of rotation, by a dragmember. This allows the driven member to overrun the driving memberwithout engaging the rollers bearings on the cams. However, when thedriving member begins to overrun the driven member, the rollers quicklyengage and torque from the driving member is transfered to the drivenmember.

Drag members add inefficiency to the roller clutch assembly because thedrag forces reduce fuel efficiency and generate substantial heat,particularly during high speed operation. That heat requires the use ofspecific materials that can withstand the friction and high temperaturesfor long periods of time. Typically, such materials are more expensivethan less wear-resistant and heat-resistant materials.

Sometimes the driven member can overrun sufficiently to overcome theforce that retards the rollers. The rollers then advance, that isprogress forward in the same direction of rotation as the races. If therollers advance sufficiently, they can engage the cams so that thedriven member actually drives the other (defeating the overrun feature).If this occurs, the clutch can lock up and remain in the engagedconfiguration even when the driven member is not trying to overrun thedriving member. This condition causes an awkward driving experience.

SUMMARY OF THE INVENTION

The transfer case of the present invention is provided with high and lowranges of operation for different environmental conditions. A neutralposition is also provided wherein the output shafts are not engaged tothe input shaft. The transfer case is shifted between these positions(high-neutral-low) by an actuator assembly that includes a compliantshift fork that is moved by a variable speed shift rail. The shift railis driven by a pin drive mechanism that is rotated at a substantiallyconstant velocity by a motor. Due to the relationship between the pindrive and the shift rail, the constant rotational velocity of the pindrive moves the shift rail at variable speeds, faster between positionsand slower proximate to the positions. This variable speed drive allowsmore precise positioning of the shift fork and less complicated positionsensors and motor controllers.

Various advantages and features of novelty which characterize theinvention are particularized in the claims forming a part hereof.However, for a better understanding of the invention and its advantages,refer to the drawings and the accompanying description in whichpreferred embodiments of the invention are illustrated and described.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a cross-sectional view of a preferred embodiment of a transfercase of the present invention in a first operating condition.

FIG. 2 is an enlarged cross-sectional view of a portion of the transfercase of FIG. 1 in a second operating condition.

FIGS. 3a and 3b are enlarged cross sectional views taken along line 3--3of FIG. 2 showing two embodiments of a latch.

FIG. 4 is an enlarged detailed view of a cross section of a modulateddrag shoe of a preferred embodiment of an overrunning roller clutch ofthe transfer case of FIG. 1.

FIG. 4a is a partial elevation view of a drag shoe and grounding memberas seen from line A--A in FIG. 4.

FIG. 5 is an enlarged cross-sectional view taken on line 5--5 of FIG. 2.

FIGS. 6a, 6b, and 6c show detailed cross-sectional views of an actuatorand pin drive mechanism as viewed along line 6--6 of FIG. 2.

FIG. 7 is a perspective view of a friction ground of the presentinvention.

FIG. 8 is a perspective view of a roller bearing cage of the presentinvention.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS

Vehicles having multiple drive axles, e.g., 4-wheel drive vehicles, needto divide torque from a common power source to front and rear axles ofthe vehicle. The division, or allocation, of torque between the frontand rear wheels from the power source is accomplished by a transfercase.

FIG. 1 shows an exemplary embodiment of the present invention. Atransfer case 10 has a housing 12 that covers a gear reduction assembly16, a coupling mechanism 18, a roller clutch assembly 20, and anactuator mechanism 22. An input shaft 24 receives power from a powersource, such as a motor-driven transmission, and delivers torque intothe transfer case. A first output shaft 26, substantially in line withthe input shaft 24, is selectively coupled to the input shaft 24 or thegear reduction assembly 16 by the coupler mechanism 18. In a preferredembodiment, the first output shaft 26 will transmit torque to a rearaxle of a vehicle.

The roller clutch 20 selectively couples the first output shaft to asecond output shaft 28. A sprocket 30 is formed in a portion of theroller clutch 20 and is coupled through a drive chain 32 to a secondsprocket 34 to complete power transfer to the second output shaft 28.

A ball bearing assembly 29a and roller bearing assembly 31 support theinput shaft 24. The first output shaft 26 is supported by ball bearingassembly 29b adjacent the output end 26a of shaft 26 and the rollerbearing 31 assembly adjacent the input end 26b. A ball bearing assembly29c and needle bearing assembly 33 support the second output shaft 28.Input shaft 24 and output shaft 26 are rotatable either independently orin conjunction with each other.

The input shaft 24 receives torque from a power source (not shown) androtates within the housing 12 and drives the gear reduction assembly 16.The coupler mechanism 18 is always coupled to the first output shaft 26and selectively couples to the input shaft 24 or to the gear reductionassembly 16 or idles therebetween in neutral. When the first outputshaft 26 is coupled to the input shaft 24, the transfer case is in highrange. When the first output shaft 26 is coupled to the gear reductionassembly 16, the transfer case is in low range. The transfer case is inneutral when the coupling mechanism is connected to neither the inputshaft nor the gear reduction assembly. The coupling mechanism 18 isshifted by the actuator assembly 22.

When the roller clutch 20 is engaged, the first output shaft 26 drivesthe second output shaft 28 via the chain-sprocket combination 30, 32,34.

The components of the transfer case 10 are explained in greater detailbelow.

Gear Reduction Assembly

In its preferred embodiment, the gear reduction assembly 16 comprises ahelical planetary gear assembly 36, which includes a sun gear 38, aplurality of planet gears 40, and a fixed ring gear 42. The planet gears40 are rotatably supported on a floating carrier 44. In the exemplaryembodiment, the planetary gear 36 includes four planet gears 40. Thehelical sun gear 38 is formed on, or fixedly secured to, the input shaft24. The sun gear 38 is meshingly engaged with helical planet gears 40,which, in turn, are meshingly engaged with the fixed, helical ring gear42. Rotation of the sun gear 38 rotates the planet gears 40 which engagethe ring gear 42. Because the ring gear is fixed, the planet gears walkaround the inside of the ring gear thus moving the carrier 44 so thatthe carrier rotates about the axis of rotation of the input shaft.

Thrust plates (or rings) 46, 48 prevent the carrier from moving axiallyunder loads created by driving the helically threaded gears. The thrustplate 46 is annular and is coupled to the housing 12 by a retaining ring46a and presses against the fixed ring gear 42. Structurally, the thrustplate 46 includes substantially flat outer and inner disk portions 46band 46c respectively with a conical portion 46d intermediate the diskportions. An inner margin of the thrust plate 46 is coupled to alow-friction ring 46e that presses against the carrier 44. The thrustplate 48 is annular and is positioned directly between the carrier andthe housing. Thus the thrust plates 46 and 48 retain the carrier 44without substantially inhibiting the carrier's rotation. Preferably thethrust plate is fabricated of steel and the low-friction rings are aself-lubricating bearing material, such as nylon.

The carrier 44 includes an internally splined extension 50 with splineteeth 50a for connection to the coupling mechanism 18 as describedbelow. When the input shaft 24 drives the planet gears 40, and hence thecarrier 44, the carrier rotates at a lower rotational velocity than theinput shaft. Preferably, the gear ratio is in a range of approximately2.5 to 1 to 3.0 to 1.

The input shaft 24 also includes an externally splined portion 52 thatcan directly couple to the coupling mechanism 18 for directly drivingthe output shaft 26. Because the splined portion 52 is part of the inputshaft 24, it rotates at the same speed as the input shaft.

Accordingly, when the output shaft 26 is coupled to the gear reductionassembly 16, the output shaft 26 rotates at a slower rotational velocitythan the input shaft 24. However, when the output shaft 26 is coupled tothe input shaft 24 (at its splined portion 52), the input shaft andoutput shaft rotate at the same speed.

Coupling Mechanism

The coupling mechanism 18 is shiftable to couple the first output shaft26 either to the input shaft 24 (high range) or the gear reductionassembly 16 (low range), or to decouple the first output shaft from both(neutral). The coupling mechanism comprises a collar 56 having aradially extending flange 58. The collar 56 includes an internal splinedsurface 60 that is received on an externally splined portion 64 of thefirst output shaft 26. Because the connection between the collar and thefirst output shaft is a splined connection, the collar is free totranslate axially along the first output shaft, but rotates with thefirst output shaft. The internal splined surface 60 also is configuredto be received in a splined engaging condition on the splined portion 52of the input shaft 24.

The collar 56 also includes an externally splined portion 62 adapted,when appropriately positioned, to couple to the splined carrierextension 50.

In FIG. 1, collar 56 is illustrated shifted to the left to engage theinput shaft's splined portion 52 and the output shaft's splined portion64, thus directly coupling the input shaft 24 to the output shaft. Thisis the high range drive mode.

In FIG. 2, the collar 56 is shown shifted to the right in the drawingwhereby splined portion 62 engages the splined carrier extension 50.Accordingly, the output shaft 26 is coupled to the input shaft 24 viathe gear reduction assembly 16. In this mode, the output shaft rotatesat the speed of carrier 44, and thus slower than the input shaft. Thisis low range mode.

The coupling mechanism 18 also has a neutral position (shown partiallyin dashed outline in FIG. 2) in which the collar 56 is not engaged witheither the carrier spline extension 50, or the spline connection 52 ofthe input shaft 24. When in neutral, neither output shaft 26, 28receives torque from the power source.

Alternative configurations exist for the gear reduction assembly and thecoupling mechanism. For example, spur gears may be used in the planetarygear or the coupling device could interface with the input shaft or gearassembly differently.

Roller Clutch

The roller clutch assembly 20 includes an inner race 70 that is formedon, or affixed to, a portion of a surface of the first output shaft 26.As the driving member of the roller clutch, the inner race 70 has aplurality of cam surfaces 71 (FIG. 3b) for engaging a respective numberof rollers, or needle rollers, 72. Preferably, the cam surfaces areflat, but other configurations may also be suitable.

The rollers are maintained in position by a roller cage 74 (FIGS. 1, 2,3 and 8) that extends circumferentially around the inner race 70 andextends axially outward, forming a skirt 76 having an end ring 77. Withreference to FIG. 8, the roller cage further includes roller openings178 separated by tangs 108. At one end of the openings 178 and tangs 108is an index ring 180 that comprises beveled keyways 182 havingsemi-circular keys 184 terminated by beveled ends 186 and stop ends 188.Alternately to the beveled ends 186, the keyway 184 could have stops ateach end and be sized sufficiently to accommodate movement of the latchas described below.

Preferably, the cage 74 also includes shoe retention legs 190 and shoeseparators 192 for retaining and positioning drag shoes 90 as describedbelow. Respective margins 194, 196 of the legs and separators arebeveled to assist with shoe retention and positioning.

The cage may further be provided with an annular notch 198 for receivinga drag ring 84, described below.

An outer race 78 is formed along an interior surface of an extension 80of the sprocket 30. The sprocket 30 is journaled for rotation on thefirst output shaft 26 and held in position by a retention ring 82.Preferably, the outer race 78, in this embodiment, is substantiallycylindrical, having no cam surfaces thereon.

Located radially inward of the outer race 78, and in contact with theouter race and the cage 74, is a drag ring 84. Preferably, the drag ring84 is a resilient ring located on the roller cage 74 and in contact withthe outer race 78.

Located adjacent, but outside, the races 70, 78 is an annular frictionground 86 that is journaled on the first output shaft 26. The frictionground includes a plurality of circumferentially located pockets 88 forreceiving a locking device 112, as identified below. A suitable materialfor the friction ground is powdered metal (steel).

Located radially outward of the friction ground 86 are a plurality ofdrag shoes 90 that can press against, and ride on, the friction ground86. The shoes are coupled to the skirt 76 to be held generally in placeand so that drag forces on the drag shoes are transferred to the rollers72 by the cage. The drag shoes are substantially semi-circular andextend through an approximate quarter circle arc. Preferably, the dragshoes 90 are provided with a saddle 91 that is positioned radiallyinward of a respective shoe retainer leg 190 (FIG. 4b). Each shoeterminates proximate a respective shoe separator 192. The shoes 90 maybe provided with beveled ends that are compatible with the beveledmargins 196 of the separators 192.

In cross-section profile (FIG. 4), the drag shoes have a drag surface200 and upper surface 202. The shoe drag surface 200 slides on the aground member 86 as described below. The upper surface 202 is positionedradially inward of cage skirt 76. Preferably, there are four shoeslocated circumferentially about the friction ground 86. A suitablematerial for the drag shoes 90 is carbon-filled polyphenylene sulfidewith PTFE.

A garter spring 92 is located about the drag shoes 90 thus urging themradially inwardly against the drag surface 96 of the friction ground 86.As best noted in the detail of FIG. 4, there is a gap 94 between thedrag shoes 90 and the skirt 76. This gap allows the drag shoes 90 to bemoved away from the friction ground 86 at high RPM, but not to bereleased totally from the system.

When the friction ground 86 is grounded, that is, held stationary withrespect to the housing 12, an annular outer surface 96 of the frictionground provides a drag surface for the shoes 90. When the shoes ride onthe drag surface 96 a drag force is created that is transfered to theroller cage 74. When the first output shaft 26 rotates (and hence theroller clutch and drag shoes 90 are rotated) at a sufficient rotationalspeed, centrifugal force urges the drag shoes radially outwardly againstthe radially inwardly directed biasing force of the spring 92. At apreselected speed, determined primarily by the mass of the shoes and theforce of the spring 92, the shoes are urged radially outward away fromthe drag surface 96 thus reducing or eliminating the drag force. The gap94 allows such movement.

In a preferred embodiment the drag shoes 90 and garter spring 92 aredesigned so that the drag shoes unload from the friction ground atapproximately 80 miles/hour. It has been determined that unloading canoccur at this speed by selecting the following approximate parameters:Four, quarter-circle drag shoes 90 weighing 0.0126 pounds each andhaving a drag surface that is 0.668 square inches. A coefficient offriction between the drag shoes 90 and the friction ground isapproximately 0.11. The garter spring 92 has a free length of 8.553inches, an installed length of 13.512 inches and an initial tension of0.736 pounds. The friction ground has an outside diameter of 3.467inches. These parameters are suitable in a system having an axle ratioof approximately 3.73 to 1.00. As the shoes wear, it is expected thatthe speed at which the shoes unload from the friction ground willincrease by five to eight miles per hour.

The operation of the variable force drag shoes is explained in greaterdetail below.

In a preferred embodiment of the roller clutch assembly 20 of thepresent invention, there is a latch 100 (FIGS. 2 and 3a) that is indexedto the inner race 70 by means of a flat 102 (alternatively keyways couldbe used) so that the latch 100 rotates with the first output shaft 26.The latch 100 includes a latch plate 103 having loosely coupled latcharms 104 that can move radially outward under centrifugal force when theoutput shaft 26 is rotating at sufficient speed.

The latch arms 104 are held on the plate 103 by retention wings 105located at ends 106 of the latch arms. The latch arms 104 are free tomove along the plane of the plate 103, but are prevented from leavingthe plane of the plate by the retention wings 105. Retention springs 107are coupled to the retention wings 105 and press against the latch arms104 to bias the arms radially inward and to guide the motion of the armsinto the desired radial motion when the latch is subjected tocentrifugal force. The latch arms 104 further include tabs 109 that areconfigured to engage the semi-circular keyway 180 of the roller cage 74.The operation of the latch 100 is described below.

In an alternative embodiment shown in FIG. 3b, the latch 200 includesthin, flexible arms 202 that extend circumferentially part-way aroundthe perimeter of the latch. Fingers 204 are located at distal ends ofthe arms. The fingers 204 are sized to fit within the keyways 180. Thearms 202 are flexible so that when the latch 200 rotates at sufficientvelocity, the arms 202 are urged outward and the fingers 204 engage thekeyways 180, as shown in dotted lines in FIG. 3b at 202' and 204'.

When the latch 100 (or 200) is engaged, the cage 74 is coupled to theinner race 70 (and thus to the first output shaft 26). The purpose andeffect of this latch mechanism will be described below in regard to theoperation of the transfer case.

Alternative latch designs are contemplated. For example, the latch couldcomprise spokes that slide radially outward under centrifugal forces.

A spacer 110 (FIG. 2) is mounted between the bearing 29b and thefriction ground 86 to act as a thrust race against axial movement of thefriction ground 86 and the roller clutch assembly 20.

The friction ground may be selectively grounded. That is, the frictionground 86 may be grounded so that it does not rotate relative to thehousing 12, or it may be ungrounded so that it rotates with the outputshaft 26. When grounded, the friction ground 86 provides the stationarydrag surface 96 upon which the shoes 90 drag to bias the roller cage androllers 72. When ungrounded, the friction ground 86 rotates with thefirst output shaft 26 and therefore does not bias the rollers in aretarding direction.

A locking hook 112 (FIGS. 1 and 2) is shiftable axially within thehousing to engage or disengage the friction ground 86. In FIG. 1, hook112 is shown engaging a pocket 88 of the friction ground 86 to groundthe friction ground so that it does not rotate relative to the housing12. In FIG. 2, the locking hook 112 is shown disengaged, or withdrawn,from the pockets 88 of the friction ground, thus ungrounding thefriction ground and allowing it to rotate relative to housing 12. Thelocking hook 112 is controlled by the position of the actuator mechanism22 as will be described in greater detail below.

Actuator Mechanism

As noted above, the coupling mechanism 18 moves, or is shiftable,between (a) high range, (b) low range, and (c) a neutral position toengage to the first output shaft 26 (a) directly to the input shaft, (b)to the planetary gear carrier 44, or (c) to an idle positiontherebetween, respectively. Movement of the coupling mechanism collar 56is accomplished by the actuator mechanism 22.

The actuator mechanism includes a power source 122, such as an electricmotor, and a planetary gear drive 124 (FIG. 5). The power source 122 isarranged to drive a sun gear 126 that meshingly engages a plurality ofplanet gears 128 that are journaled on a floating carrier 130. Inmeshing engagement with the planet gears 128 are outer annular rings132a and 132b. The annular ring 132b is fixed to the power source 122(or a housing thereof). The annular ring 132a is floating, that is notfixed, and is coupled to a pin drive device 138 having, preferably, twodrive pins 140a and 140b (collectively 140) that terminate in a cap thatis supported by a bearing 142 within the housing (FIG. 2). The annularrings 132a and 132b are provided with teeth along an inner surfacethereof for engaging the planet gears 128. However the annular rings132a and 132b have a different number of teeth so that as the planetgears and carrier 130 walk around the inside of the annular rings 132aand 132b, the floating annular ring 132a is forced to rotate thus movingthe pin drive device.

The drive pins 140a and 140b engage notches 144 formed in an elongateshift rail 146 that is slidingly supported in the housing 12.Accordingly, as the power source drives the planetary gear 124, the pindrive device 138 rotates about an axis of rotation 148 and the drivepins 140a and 140b translate about the axis of rotation, as can best beseen in FIGS. 6a-6c. The rotation of the pin drive device 138 engagesthe shift rail 146 and causes it to translate along its longitudinalaxis.

Although not shown, a preferred embodiment of the actuator assembly 22includes a rotational position sensor to monitor the drive pin devicerotation and deenergize the motor when the actuator reaches desiredpositions.

The shift rail is compliantly coupled to a shift fork 150 by means of acompliance spring 152. The shift fork extends upward and couples to theflange 58 of collar 56 so that translation movement of the shift fork150 likewise causes the collar 56 to translate into the variouspositions corresponding with the high range, low range, and neutralpositions. The compliant coupling between the shift rail 146 and theshift fork 150 produced by spring 152 permits relative movement betweenthe shift rail and the shift fork when the collar is not able to slidefreely. For example, if the collar 56 is blocked from sliding when theshift rail 146 moves, the compression spring 152 will be compressed andprovide a force against the shift fork 150 which, in turn, will providea force against the collar 56, urging it to move. When the splines(e.g., spline portions 52 and 60 for example) are properly aligned thespring force will cause the collar to move.

The shift fork 150 also connects to the locking hook 112 at a hookreceptacle 154. With reference to the orientation of FIG. 2, it can beseen that movement of the shift fork 150 to the right will push againstthe locking hook 112, moving it out of engagement with the frictionground 86. Conversely, when the shift fork 150 moves to the left, abiasing spring 156 urges the locking hook 112 to move to the left and,when the hook is properly aligned with a pocket 88, it will enter thepocket, thereby grounding the friction ground member 86.

Operation of the Transfer Case

With particular reference to FIGS. 1 and 2 and other figures as noted,the operation of the transfer case 10 of the present invention will beexplained. First, high range operation is discussed.

High Range Mode

The basic purpose of the transfer case is to receive torque at the inputshaft 24 and distribute torque between the front and rear axles basedupon conditions to which the vehicle is subjected. Under simpleoperating conditions when the transfer case is in high range mode (asshown in FIG. 1) and the vehicle is traveling straight and at moderatespeed along a level, high-friction surface (e.g. asphalt) the drive lineof the vehicle will proceed directly from the input shaft 24 to thefirst output shaft 26. Under these conditions, the road surface rotatesthe front wheels at substantially the same rate of speed as the rearwheels due solely to their contact with the ground surface. The secondoutput shaft 28 is in direct drive relationship with the front wheels.

Various vehicle and environmental conditions affect the drive path. Forexample, in a front-engine vehicle, the rear wheels may slip on the roadsurface more than the front wheels, even on normal, level road surfaces.Thus, the inner race 70 located on the first output shaft 26 willattempt to rotate faster than the outer race 78 thereby causing theroller clutch to engage. Under these circumstances, the torque isdirected primarily to the front axle.

In a rear-engine vehicle, or a vehicle towing a trailer, the rear wheelsmay not slip more than the front. Without rear wheel slip, the innerrace generally would not attempt to rotate faster than the outer raceand the clutch normally would not engage. Thus, the torque would beprimarily directed to the rear axle.

When the vehicle turns through a corner, the front wheels can rotatefaster than the rear wheels thus causing the outer race 78 to overrunthe inner race 70. Conversely, if the rear wheels slip due to a lowfriction surface, or other cause, the clutch 20 will engage and transferinput torque to the front axle.

In high range mode, the internal splines of collar 56 directly engagesplines 52 and 64 on shafts 24, 26, respectively (FIG. 1). Thus, inputtorque delivered to the input shaft 24 is transferred directly to thefirst output shaft 26 by virtue of the direct connection. When thecoupling mechanism is in the high-range position (FIG. 1), the biasingspring 156 urges the locking hook 112 into one of the pockets 88 of thefriction ground 86, grounding it. Because the friction ground 86 isjournaled on the first output shaft 26, it remains stationary while thefirst output shaft 26 rotates within it.

The drag shoes 90 rest upon the drag surface 96 of the friction ground86 and rotate with the rollers and roller cage. As the roller clutchassembly rotates, the drag shoes 90 drag across the drag surface 96. Thegarter spring 92 provides an inwardly directed radial force thus urgingthe shoes 90 against the drag surface 96 (FIG. 4).

The operation of the system is speed dependent. As the rotationalvelocity of the first output shaft 26 increases, the system changes thuscreating different dynamics between the system components.

When the first output shaft 26 rotates at a low rotational velocity, thedrag shoes 90 rest upon, and drag across the stationary friction ground86.

Assuming that the vehicle is traveling straight along a flat road havinggood surface friction, the inner race 70 and outer race 78 are rotatingat substantially the same rotational velocity. The drag shoes 90 dragacross the drag surface 96 of the grounded friction ground 86 creating adrag force that acts on the cage 74 to retard the rollers so that theouter race 78 is free to overrun the inner race at any time. It isparticularly important that the outer race 78 be able to overrun theinner race at lower speeds as this is when the sharpest turning occurs.Additionally, with the rollers retarded, they are in a position toquickly engage in the event that the inner race 70 rotates faster thanthe outer race as is common when an operator attempts to move thevehicle from a low friction position, such as when the rear wheels arelocated on ice or sand. Accordingly, when attempting to move forwardfrom a stopped position and the rear wheels slip, the roller clutch willengage very quickly with very little relative movement between the innerand outer races and thus quickly transfer torque to the front axle.

As the vehicle moves faster and the first and second output shafts 26,28 begin to achieve a higher rotational velocity, the drag shoes 90likewise rotate faster and a centrifugal force acts radially outwardlyon the shoes in opposition to the force of the garter spring 92. Asnoted, there is a gap 94 between the upper surface 202 of the shoes 90and the roller cage skirt 76 thus providing space into which the shoescan move. As the rotational velocity increases, the centrifugal forceincreases proportionally and the shoes start to lift off of the frictionsurface 96 thus reducing the amount of drag force that acts on the cage74 and rollers 72. The amount of force on the rollers is less necessaryat higher speeds because the vehicle is substantially less likely tomake sharp turns and rear wheel slip is not as likely once the vehicleis moving at a substantial speed.

Reducing the friction force of the drag shoes at high speed isbeneficial because less heat is generated between the surfaces and lessfrictional wear occurs. A greater range of materials may be used for thefriction ground 86 and the shoes 90 due to the reduction in frictionalwear and heat produced. This promotes longer useful life for theoperating components in this part of the system.

The factors that influence the separation of the shoes from the frictionground include the spring rate of the garter spring 92, the initialspring tension, the mass of the shoes 90, the radius of the shoes 90from the center of rotation, and the speed of rotation. In the currentembodiment, the shoes 90 are designed to separate completely from thefriction ground 86 at approximately 80 miles per hour (plus or minus 10miles per hour). At typical highway speeds, the shoes 90 will have begunto lift away from the friction ground 86 and the force retarding therollers due to the drag shoes 90 will be reduced.

The drag ring 84 moves with the outer race 78 and urges the roller cage74 to advance. When the drag shoes 90 are frictionally sliding onfriction ground 86 the drag force so generated by the drag shoes farexceeds the force of the drag ring on the roller cage. Thus, the dragring does not cause the rollers to advance when the drag shoes areengaged with the drag surface 96. However, when the shoes are off thedrag surface and the outer race rotates faster than the inner race(front wheels moving faster than the rear wheels) then the drag ring 84will urge the rollers to advance toward their advanced engagementposition thereby engaging the roller clutch.

When lock-up occurs in previously known roller clutches, the rollerclutch assembly could remain engaged even when the vehicle returned tolower speeds. Such lock-up can be particularly noticeable when a vehicleis traveling along a highway and the roller clutch engages in theadvanced engagement position and thereafter the vehicle exits thehighway onto a down ramp, stops and makes a sharp turn onto a side road.This condition can be disconcerting to the vehicle operator if theroller clutch remains engaged. The latch 100 will prevent such lock-up.

Mounted on the first output shaft 26, the latch 100 rotates with theshaft. At lower velocities, the latch arms 104 are retracted by thesprings 107 into a nonengaged configuration as is shown in solid outlinein FIG. 3a. As the velocity of the first output shaft increases, thearms 104 are urged outward due to the centrifugal force created by therotation of the shaft. With sufficient centrifugal force exerted, thearms move far enough that the tabs 109 will engage the cage 74 at thekeyways 180 (as shown in dashed line 111 for one of the arms) thuscoupling the roller cage 74 to the first output shaft 26 so that thecage rotates with the output shaft 26. The keyways 180 are arrangedrelative to the roller openings 178 so that the latch arms 104 engagethe cage 74 when the rollers are approximately halfway between theretarded engaged position and the advanced engaged position The trailingend of each keyway 180 has the beveled end 186 so that if the inner racebegins to overrun the outer, the tabs 109 will encounter the beveledends and easily slip past the keyway to allow the cage and rollers toretard so the clutch can lock. Thus, the beveled end permits the clutchto engage in the retarded position, but prevent engagement in theadvanced position.

Neutral

The coupling mechanism 18 has an intermediate position between high andlow range, shown partially in dashed line in FIG. 2, in which the firstoutput shaft 26 is not engaged to either the input shaft 24 or to thegear reduction assembly 16. In this position the locking hook 112 isengaged with the friction ground to ensure that the drag shoes 90 biasthe rollers so that the clutch assembly 20 continues to operate as anoverrunning clutch. This position is particularly useful for situationsin which the vehicle is being towed.

Low Range Mode

When the coupling mechanism 18 is moved into low range as shown in solidline in FIG. 2, collar 56 connects the spline portion 64 of the firstoutput shaft 26 with the carrier 44 of the planetary gear 36. Thiscauses the first output shaft 26 to rotate slower than the input shaft24. As the actuator assembly 22 moves the collar 56 it also forces thelocking hook 112 out of engagement with the friction ground 86. Thefriction ground 86 remains frictionally coupled to the drag shoes 90 andis now free to rotate with the output shaft 26 and the roller clutchassembly 20. When the friction ground is thus ungrounded, the dragsurface 96 rotates and the drag shoes 90 provide no drag force to retardthe roller cage and rollers. Thus the clutch 20 has no substantialoverrunning function during low range operation.

It has been found beneficial to have the vehicle in full-time four wheeldrive mode when the vehicle is in low range and operating on rough orlose terrain. Without the biasing force on the rollers, all relevantmovement between the inner race 70 and outer race 78 will cause theroller clutch to engage so that forces on the faster axle drive theslower axle. This is particularly beneficial when, for example, thevehicle is in low range mode and going forward down a steep embankmenthaving a lose ground surface. The front wheels can provide compressionbraking to the rear wheels when the roller clutch is thus engaged.

The drag ring 84 is further effective in this configuration because ittends to advance the rollers into engagement when the outer race 78 ismoving faster then the inner race. Without the drag ring 84 there may besituations in which the rollers could be retarded so that the frontwheels would overrun the rear wheels and not provide any compressionbraking.

Actuator Operation

As mentioned, the actuator assembly 22 provides the force to move thecollar 56 of the coupling mechanism 18 between high range, low range andneutral. The actuator motor 122 (FIG. 5) rotates the sun gear 126 of theplanetary gear 124. Rotation of the sun gear 126 rotates planet gears128 which thus cause the floating annular ring 132a to rotate. Attachedto the annular ring is the pin drive device 138 having the drive pins140. As can best be seen in FIGS. 6a-6c, the drive pins 140 engagenotches 144 in the shift rail 146. Because the pin drive device 138 isfixed along its center of rotation, rotating the pin drive device 138will cause a drive pin 140 to engage the notch and force the shift raillongitudinally to the left or right (depending upon the direction ofrotation).

In FIG. 6a, the shift rail 146 and pin drive device 138 are shown in theposition which would cause the coupling mechanism 118 to be in the highrange position. Thereafter, rotating the pin drive device 138 in theclockwise direction (as viewed from FIGS. 6a-6c) causes a drive pin 140to engage a notch 144 to force the shift rail longitudinally to theright.

For the following explanation of the operation, it is necessary to setup references axes. Accordingly, a drive pin axis 160 extendstransversely through the centers of drive pins 140a and 140b and centralaxis 148. A longitudinal axis 162 is defined extending longitudinallythrough the shift rail 146. An angle is formed between the drive pinaxis 160 and the longitudinal axis 162, defined as phi (φ).

As noted, the configuration of the pin drive device 138 and shift rail146 as shown in FIG. 6a represent the actuator when the couplingmechanism is in the high range position. In this configuration, φ=0. Asthe pin drive device 138 begins to rotate clockwise, φ likewiseincreases and drive pin 140a slowly begins to move the shift rail. Whenthe pin drive device 138 gets to the position as shown in 6b whereφ=90°, the shift rail is being moved more quickly even while the drivepin device 138 maintains a constant rotational velocity. As the pindrive device rotates another 90° and reaches the position as shown inFIG. 6c, φ again approaches 0° and the actuator mechanism is in theposition corresponding to the neutral position for the couplingmechanism 18. When φ is approximately 0°, rotating the pin drive device138 produces little motion of the shift rail 146. When φ isapproximately 90°, an identical amount of rotation of the pin drivedevice 138 produces substantially greater longitudinal motion of theshift rail 146. Accordingly, it can be seen that the relationshipbetween the speed of the shift rail 146 is proportionate to the sine ofφ. The shift rail velocity is equal to ω·R·sin φ, where ω is therotational velocity of the pin drive device 138, and R is the effectiveradius of the drive pins 140 from the center of rotation 148 of the pindrive device.

This relationship is beneficial for several reasons. When the actuatoris moving and the coupling device 18 is approaching one of the threepositions (high, low or neutral), φ will approach zero. Accordingly, themovement of the shift rail will not as be sensitive to slight motions ofthe pin drive device 138, thus making it easier to accurately sense andstop the actuator assembly 22 when the coupling mechanism has reachedthe desired position.

Conversely, system speed is greatest when φ is approximately 90°. Thus,the actuator mechanism is able to move the shift rail 146 quicklybetween positions but naturally slows as φ reaches zero, whichcorresponds to the shift rail reaching a desired position. The positionrepresented in FIG. 6c represents the neutral position of the couplingmechanism 18. Rotating the pin drive device 138 another 180° clockwisefrom FIG. 6c would correspond with moving the coupling of mechanism intothe low range position.

Numerous characteristics and advantages of the invention have been setforth in the foregoing description, together with the details of thestructure and function of the invention. The novel features hereof arepointed out in the appended claims. The disclosure is illustrative only,and changes may be made in detail, especially in matters of shape, sizeand arrangement of parts within the principle of the invention to thefull extent indicated by the broad general meaning of the terms in theclaims.

What is claimed is:
 1. A transfer case, comprising:(a) a housing: (b) aninput shaft rotatably coupled to the housing; (c) an output shaftrotatably coupled to the housing; (d) a torque transfer mechanismoperable to impart first and second torque transfer conditions betweenthe input and output shafts; (e) a movable shifter mechanism operable toselectively shift between said first and second torque transferconditions; and (f) and an actuator for moving the shifter mechanism,the actuator including a power source, a pin drive device defining anaxis of rotation and having at least two elongate drive pins extendingsubstantially parallel to and arranged to translate circumferentiallyabout the axis of rotation as the pin drive device rotates, and a shiftrail having a plurality of notches positioned to be engaged by the drivepins, whereby the power source rotates the pin drive device about theaxis of rotation, the drive pins translate circumferentially about theaxis of rotation and at least one drive pin engages at least one notchand urges the shift rail to move longitudinally thereby moving theshifter mechanism.
 2. The transfer case of claim 1 wherein a lineextending between the two drive pins defines a pin axis that extendssubstantially normal to the elongate drive pins and passes through acenter of the drive pins, the shift rail has a longitudinal axis, thepin axis and the longitudinal axis define an angle of incidencetherebetween, and the velocity of the shift rail is proportionate tospeed of rotation of the drive pins and the sine of the angle ofincidence.
 3. The transfer case of claim 1 wherein the power source is amotor having a substantially constant rotational velocity, a lineextending between the two drive pins defines a pin axis that extendssubstantially normal to the elongate drive pins and passes through acenter of the two drive pins, the shift rail defines a longitudinalaxis, the velocity of the shift rail is greatest when the pin axis issubstantially normal to the longitudinal axis and the velocity of theshift rail is least when the pin axis is substantially parallel to thelongitudinal axis.
 4. The transfer case of claim 1 wherein the shiftermechanism has first and second positions, the first position couples thefirst output shaft to the torque transfer mechanism and the secondposition decouples the first output shaft from the torque transfermechanism, and a line extending between the two drive pins defines a pinaxis that is substantially normal to the elongate drive pins and passesthrough a center of the two drive pins, the shift rail has alongitudinal axis, the pin axis and the longitudinal axis define anangle of incidence therebetween, and when the shifter mechanism is inthe first and second positions the angle of incidence is substantiallyzero.
 5. The transfer case of claim 1 wherein the shifter mechanism hasfirst, second, and third positions, the first position couples the firstoutput shaft to the torque transfer mechanism, the second positiondecouples the first output shaft from the torque transfer mechanism, andthe third position couples the first output shaft to the input shaft,and two of the drive pins define a pin axis that is substantiallyorthogonal to the elongate drive pins and passes through a center of thetwo drive pins, and the shift rail defines a longitudinal axis, and thepin axis and the longitudinal axis define an angle of incidencetherebetween, and the when the shifter mechanism is in the first,second, and third positions the angle of incidence is substantially zerodegrees.
 6. The transfer case of claim 1 wherein the shift rail has fournotches defined therein for engaging the drive pins.
 7. The transfercase of claim 1 further comprising a shift fork compliantly coupled tothe shift rail and coupled to the shifter mechanism, whereinlongitudinal motion of the shift rail urges the shift fork to move whichin turn moves the shifter mechanism.
 8. The transfer case of claim 1further comprising a selectively grounded friction collar, at least onefriction shoe that drags against the collar, and a locking device thatengages and grounds the friction collar, wherein the locking device iscoupled to the actuator so that motion associated with the actuatorengages and disengages the locking device and the friction collar. 9.The transfer case of claim 1, which further comprises a selectivelygroundable member which may be either rotatable or grounded againstrotation, a locking device shiftable between a first position engagingand grounding said grounding member and a second position permittingrotation of the groundable member, and the locking device is coupled tothe actuator so that motion of the actuator shifts the locking devicebetween said first and second positions.
 10. The transfer case of claim1, which further comprises a compliant coupling between the shiftermechanism and the actuator so that operation of actuator urges theshifter mechanism to move when movement of the shifter mechanism isunobstructed.
 11. The transfer case of claim 1 further comprising aroller clutch, and a second output shaft coupled to the first outputshaft through the roller clutch, the roller clutch including a biasingdevice ground member and a locking device that engages the ground memberto ground it and disengages the ground member to unground it, thelocking device being coupled to the actuator wherein movement of theactuator moves the locking device thereby engaging and disengaging thelocking device with the ground member.
 12. A transfer case for avehicle, comprising:(a) a housing; (b) an input shaft rotatably coupledto the housing; (c) a torque transfer mechanism coupled to the inputshaft; (d) a first output shaft; (e) a shift collar moveable between afirst position that couples the first output shaft to the input shaftand a second position that couples the first output shaft to the torquetransfer mechanism; (f) an actuator for moving the shift collar betweenthe first and second positions, the actuator including a motor, a drivepin mechanism and an elongate shift rail, whereby energization of themotor rotates the drive pin mechanism which longitudinally translatesthe shift rail to move the shift collar between the first and secondpositions.
 13. The transfer case of claim 12 further comprising acompliant spring coupling the shift rail to the shift collar so thatlongitudinal movement of the shift rail acts on the spring, the springurges the shift collar to move, and the shift collar so moves when itsmotion is unobstructed.
 14. The transfer case of claim 12 wherein thedrive pin mechanism includes a sun gear, planet gears located on acarrier and a fixed annular gear, the carrier having an axis of rotationand including a plurality of elongate pins arranged substantiallyparallel to the axis of rotation and spaced apart from the axis ofrotation so that the pins translate about the axis of rotation as thecarrier rotates, and wherein the shift rail includes a plurality ofnotches for receiving the pins so that rotation of the carrier moves thepins about the axis of rotation and the pins engage the notches therebylongitudinally moving the shift rail to urge the collar to move.
 15. Thetransfer case of claim 12 further including a shift fork coupled to theshift rail and shift collar, and movement of the shift rail urges theshift fork to move to produce movement of the shift collar.
 16. Thetransfer case of claim 12 further comprising a selectively groundablemember and a locking member, the locking member being coupled to, andmovable with, the shift rail wherein the locking member engages andsecures the groundable member in a grounded position when the shift railhas moved the collar into one of said first and second positions, andthe locking member is disengaged from the groundable member when theshift rail has moved the collar into the other of said first and secondpositions.
 17. The transfer case of claim 12 further comprising a shiftfork, a locking member, a second output shaft and a roller clutch thatselectively engages the first output shaft to the second output shaft,the roller clutch including cam elements and a selectively groundedfriction ground for biasing the cam elements, wherein the shift fork iscoupled to the shift collar and the shift rail and the locking memberand when the shift collar is in the first position the locking member isengaged with the friction ground thereby grounding the friction groundand when the shift collar is in the second position the locking memberis disengaged from the friction ground so that the friction ground isungrounded.
 18. The transfer case of claim 12 wherein the torquetransfer mechanism comprises a planetary gear assembly having a splinedconnector portion and the shift collar comprises a splined sleeve ismovable into and out of engagement with the planetary gear assembly. 19.The transfer case of claim 12 wherein the shift rail includes first,second, third and fourth notches and the gear reduction unit includes acarrier having two laterally spaced pins extending therefrom and whenthe two pins are engaged with the first and second notches the collar isin the first position and when the two pins are in the third and fourthnotches the collar is in the second position.
 20. The transfer case ofclaim 12 wherein the shift rail includes first, second, third and fourthnotches and the gear reduction unit includes a carrier with two pinsextending therefrom and when the two pins are engaged with the first andsecond notches the collar is in the first position and when the two pinsare in the third and fourth notches the collar is in the secondposition, and wherein the second and third notches have a depth into theshift rail that is greater than a depth associated with the first andfourth notches.
 21. The transfer case of claim 12 wherein the pinsextend substantially normal to the elongate shift rail.
 22. A transfercase for a vehicle drive train that transfers rotational torque from apower source to drive wheels, the transfer case comprising:(a) ahousing; (b) an input shaft adapted to receive rotational torque fromthe power source; (c) a torque transfer mechanism coupled to the inputshaft (d) a first output shaft selectively coupled to the input shaft orthe torque transfer mechanism (e) a shift actuator having at least twodrive pins that translate about a common axis of rotation, an elongateshift rail that is engaged by the drive pins and is moved longitudinallythereby when the drive pins rotate about the axis of rotation toselectively couple the first output shaft to the torque transfermechanism or the input shaft.
 23. The transfer case of claim 22 furthercomprising a shift collar that slides between a first position thatcouples the first output shaft to the input shaft and a second positionthat couples the first output shaft to the torque transfer mechanism andwherein the shift rail is coupled to the shift collar so thatlongitudinal movement of the shift rail moves the shift collar betweenthe first and second positions.
 24. The transfer case of claim 22further comprising a shift collar that slides between a first positionthat couples the first output shaft to the input shaft and a secondposition that couples the first output shaft to the torque transfermechanism, and shift fork coupled to the shift collar and the shift railand wherein longitudinal movement of the shift rail moves the shift forkthat moves and shift collar between the first and second positions. 25.The transfer case of claim 22 further comprising a shift collar thatslides between a first position that couples the first output shaft tothe input shaft and a second position that couples the first outputshaft to the torque transfer mechanism, shift fork coupled to the shiftcollar and the shift rail, and a compliance spring coupled between theshift rail and the shift fork wherein longitudinal movement of the shiftrail creates potential energy in the compliance spring that urges theshift fork and shift collar to move and the shift fork and shift collarso move when the shift collar has an unimpeded movement path.
 26. Thetransfer case of claim 22 further comprising a second output shaft and aroller clutch coupled to the first and second output shafts, the rollerclutch including first and second races and roller bearings locatedbetween the first and second races, the roller bearings having anengagement position that couples the first and second races, the rollerclutch further including at least one shoe and a selectively groundedfriction ground, the transfer case further comprising a grounding hookcoupled to the shift rail, whereby movement of the shift rail moves thegrounding hook between first and second positions that ground andunground the friction ground, respectively.
 27. The transfer case ofclaim 22 further comprising a second output shaft and a roller clutchcoupled to the first and second output shafts, the roller clutchincluding first and second races and roller bearings located between thefirst and second races, the roller bearings having an engagementposition that couples the first and second races, the roller clutchfurther including a selectively grounded friction ground and at leastone shoe that is coupled to the bearings and when the friction ground isgrounded the at least one shoe frictionally bears on the friction groundto bias the bearings out of the engagement position, the transfer casefurther comprising a grounding hook coupled to the shift rail, wherebymovement of the shift rail moves the grounding hook between first andsecond positions that ground and unground the friction ground,respectively.
 28. The transfer case of claim 22 further comprising asecond output shaft and a roller clutch coupled to the first and secondoutput shafts, the roller clutch including first and second races androller bearings located between the first and second races, the rollerbearings having an engagement position that couples the first and secondraces, the roller clutch further including a selectively groundedfriction ground and at least one shoe that is coupled to the bearingsand when the friction ground is grounded the at least one shoefrictionally bears on the friction ground to bias the bearings out ofthe engagement position, the transfer case further comprising agrounding hook coupled to the shift rail, whereby movement of the shiftrail moves the grounding hook between first and second positions thatground and unground the friction ground, respectively, and wherein thetorque transfer mechanism is a planetary gear and when the first outputshaft is coupled to the input shaft the transfer case is in high gearand the grounding hook is in the first position that grounds thefriction ground and when the first output shaft is coupled to theplanetary gear the transfer case is in low gear and the grounding hookis in the second position that ungrounds the friction ground.